Radial diffusion compressors



Sept 17, 1957 E. A. STALKER 2,806,545

RADIAL DIFFUSION COMPRESSORS Filed March 2, 1951 2 Sheets-Sheet l r 2 c C x 3| 4 B m INVENTOR.

5g 3 flaw/4 gim /C p 7, 1957 E. 4. STALKER 2,806,645

RADIAL DIFFUSION COMPRESSORS Filed March 2, 1951 2 Sheets-Sheet 2 32) E \KHKJM IN VEN TOR.

United States Patent O f 2,806,645 RADIAL DIFFUSION COMPRESSORS Edward A. Stalker, Bay City, Mich. Application March 2, 1951, Serial No. 213,597

4 Claims. (Cl. 230-120) This invention relates to axial flow compressors wherein the diffusion is accomplished in a radial direction.

An object of this invention is to provide a radial diffusion compressor of optimum proportions.

Other objects will appear from the description, drawings and claims.

This application discloses material somewhat similar to that in my application Serial No. 593,631 filed May 14, 1945, now Patent No. 2,648,492. It differs from the former application in disclosing proportions and arrangements for optimum performance of the compressor.

The above objects are accomplished by the means illustrated in the accompanying drawings in which- Fig. 1 is a fragmentary axial section through the compressor of this invention;

Fig. 2 is a fragmentary development of the rotor blad- Fig. 3 is a fragmentary development of the inlet guide vanes and rotor blading spaced apart axially to exhibit the flow vectors;

Fig. 4 is a fragmentary development of the stage blad- Fig. 5 is an axial section through a fluid diffuser;

Fig. 6 is a fragment of the rotor case shown in section;

Fig. 7 is a view of a blade and a fragment of the rotor hub in section along line 77 of Fig. 4;

Fig. 7a is a view of a blade and a fragment of an alternate form of the rotor hub in section along line 7a7a of Fig. 4; and

Fig. 8 is a fragmentary axial section of a two-stage compressor according to this invention.

Referring to the drawings and particularly to Figs. 1 and 2, the compressor is indicated generally as 10. The rotor hub 12 is mounted in case 11 for rotation with the shaft 14 in the bearings 16 and 18. The blades 31 of the rotor are carried on the hub and define a plurality of peripherally positioned passages 30. The inlet guide vanes are 32. They serve to give the fluid entering the rotor a pre-whirl in the direction of rotation, which provides an initial radial pressure to initiate diffusion in the rotor passages close to their inlets and shortens the efficient length of these passages.

The blades are parallel and, at the design condition of operation, the streamlines of flow 36 and 38 are parallel to the blades in front of and behind the rotor as shown in Fig. 2. However as shown in Fig. 1 these same streamlines diverge radially as the flow proceeds through a rotor passage. Due to the passage expansion radially the flow is diffused radially and decreases in velocity. Then according to the Bernoulli formula the static pressure will increase.

To accomplish the maximum pressure rise in the rotor, it must be operated at the highest possible tip speed which may be higher than the velocity of sound in the pumped fluid. In many compressor applications the velocity of the fluid relative to the blades should not exceed the velocity of sound in the fluid. To retain the high tip ice speed and subsonic relative fluid speed, inducer blades 32 are employed to direct the entering flow with a peripheral component in the direction of rotation as illustrated in Fig. 3. The inducer blades deflect the flow 42 to the direction 44. The velocity vector of this flow is commonly called C1. The relative flow component due to the peripheral rotation of the rotor is u. The resultant fluid velocity relative to the rotor blade 31 of the rotor 12 is V. The angle between the plane of rotation and vector C1, is )3 and the angle between the vector V and the plane of rotation is 0, the pitch angle of the blade. The angle 9 is herein called the exit angle and is defined by the tangent to the mean camber line of the inducer or stator blade section at its trailing end. The exit angle is always the smaller angle between the tangent and the plane of rotation.

The pitch angle 0 is always the smaller angle made with the plane of rotation and is positive for the blade nose directed in the direction of rotation of the rotor.

It may now be shown that there are critical angular relations between 0 and ,8 to achieve not only maximum pressure and efficiency, but to achieve even a significant compression.

The static pressure rise in passing through the rotor passages is according to the Bernoulli equation:

where p=mass density of the fluid V1=relative entering velocity C1=absolute entering velocity V2=relative leaving velocity C2=absolute leaving velocity According to trigonometry the relative velocities may If these are substituted into Equation 2 the pressure rise is expressed as P=%[u(C C cos 0] 5 An examination of this equation shows that the pressure rise will be zero for 0 equal to since the cosine then will be zero. The pressure rise will be high when 0 is substantially less than 90. For very small flow through the rotor the value of 0 could approach zero. However since the quantity of flow is always important the angle 0 must be substantial in magnitude.

At the inlet of a compressor there is static pressure drop, as follows from Bernoullis equation, to create the inflow velocity. Thus the flow through the stator or inducer vanes is accompanied by a static pressure drop.

Since the inducer vanes create velocity at the expense of static pressure, the amount of peripheral velocity supplied by the inducer vanes, should be less than the peripheral component of V in order that the static pressure rise supplied in the rotor passages by the rotor 'will be greater than that lost by the action of the inducer vanes. Thus as inFig. 3 the peripheral component a, of V is substantiall'y greater than M2 the peripheral component of. C1. This condition in the radial diffusion rotor is also satisa.

fied if the blade angle is smaller than the exit or inducer vane angle 5. Thus the blades of the rotor are to be set at a flatter exit angle than the inducer or stator vanes relative to the plane of rotation.

It is thus also clear from this'analysis that the rotor blades must have a substantial angle with respectto the axial direction to produce a compression of the fluid.

, As shown in Fig. 1 the flow from the rotor is discharged into the stator passages 70 between the blades 72.. These blades and passages are also shown in Fig. 4 representing a fragmentary development of the inlet guide vanes, the rotor blades and the stator blades.

It is known that for maximum efficiency of diffusion a tube of circular cross section such as 74 shown in Fig. 5, should have a diffusion angle D not greater than about 7 degrees or the flow will separate from the diverging walls and there will be a loss of pressure. This loss becomes excessive when the angle exceeds 12 degrees.

Some incipient or mild turbulence develops at angles above 7 degrees and at about 12 degrees the flow separates from the diverging wall and the resulting eddying flow is very inefficient and the conversion of dynamic pressure to static pressure declines. Thus about 12 degrees represents a critical angle for the diffusion with respect to pressure and efficiency.

In this invention the criticalness of the diffusion angles is recognized and means are provided to prevent this separation for very large angles D, Fig. 1, greater than 12 degrees and even of the order of 45 degrees. Thus all the diffusion in Fig. l is accomplished by the divergence being placed at the outer wall where the centrifugal force of the fluid in the rotor passages will hold the fluid to the case without separation therefrom.

It is important to use a large diffusion angle to obtain a short overall length of the compressor and light weight for a given ratio of the exit area of each rotor passage to the inlet area-or for a given ratio of radial depth at exit to the radial depth at inlet.

Fig. 7 shows a side view of a blade as seen from line 7-7, in Fig. 4. The diffusion angle of the passage is M. It should be not less than 15 degrees and preferably in the range of 15 degrees to 30 degrees. The outer edge 76 departs outward more than the inner edge 80 with respect to a reference line perpendicular to a radial line and parallel to the blade edge. The inner edge is preferably straight in the direction of the passage flow as shown in Fig. 7 and therefore shows itself as curved in the axial plane as in Fig. 1.

The inner edge of the blade near the exit should preferably not diverge toward the axis of rotation although some divergence can be tolerated. The centrifugal pressure at the inner edge of the blade is in the wrong direction to prevent flow separation from the hub. However this angle of divergence M should not exceed 12 as shown in Fig. 7a. It is measured between line 90 perpendicular to the radial direction, and the tangent 92 to the hub surface near the passage exit at the inner edge 80a of the blade shown in Fig. 7a. The reference line is also parallel to the blade surface.

To obtain high fluid pressures the ratio of exit area to inlet area should be from 1.25 to 3.0 with a preferred value in the neighborhood of 2. The ratio of exit radial depth to inlet radial depth should also be from 1.25 to 3.0 with preferred values near 2.

Also in order for the radial diffusion to give a substantial pressure rise and a high efliciency the ratio of the rotor hub diameter to the rotor tip diameter should be relatively large. Preferably the value of this ratio should be 0.7 or greater and the ratio should increase in value with successive downstream stages.

With these proportions a single stage of the compressor can produce, at subsonic speeds, pressure rises from two to three times contemporary practice for like efliciencies in axial flow machines. Since the blades are simple sheets the fabrication cost of each rotor is only about one-quarter of the cost for curved blades of airfoil section. The cost is further reduced because tests show that 6 stages for instance will produce a greater pressure ratio than 12 contemporary stages and this is accomplished in a machine of about one-half the overall length.

The axial section of the case 11 discloses a diffusion angle greater than the diffusion angle of a rotor passage since the latter is ordinarily at a substantial pitch angle to an axial plane of reference. Thus if the pitch angle of the passage is 45 degrees, the case diffusion angle will be about 1.4 times the diffusion angle of the passage where the diffusion is substantially all at the outer portion of the rotor passage.

The axial length of the rotor is of the order of the radial depth of the rotor passages.

The axial length of the stator is defined as the length of the tapered section of the case between the exit of the rotor and the exit of the stator, or the length of the stator between successive rotors in multi-rotor machine.

The stator of a radial diffusion compressor is characterized, as shown in Figs. 1 and 8, by decreasing radial depth and an axial length about equal to the maximum radial depth. The included angle E between the inner and outer walls of the stator is preferably greater than the angle M and comparable to angle D and the ratio of exit to inlet radial depth is preferably from 0.35 to 0.75.

Furthermore the exit area of the stator is comparable to and preferably smaller than the inlet area of the upstream rotor in order to achieve a satisfactory efficiency, both areas being measured in a transverse plane perpendicular to the axis of rotation. If the stator passage expanded there would be a substantial pressure rise in the stator which would cause separation of the flow from the highly curved stator blades. According to this invention the static pressure is maintained substantially constant along the stator length or actually decreased somewhat to keep the flow from separating from the curved blades.

Thus in a multi-stage radial diffusion compressor, the

case is characterized 'by alternately expanding and contracting as shown in Fig. 8. The second rotor is 102 and the second stator is 104. Since in the compressor of this invention, the fluid approaches the rotor blades parallel to them, the pumping action is very efficient as compared .to blades which curve the flow in the peripheral direction. Thus the blades of this invention do not require airfoil sections which are expensive to produce but may be made from flat stock such as sheet metal with a rounded leading edge and a tapered trailing edge. The blades preferably having leading edges :of constant radii so that they can be produced by a tool passing radially along the leading edge.

The rotors disclosed herein are axial flow rotors. That is they have blades which receive the flow transversely across the leading edges along chordwise sections of rounded leading and tapered trailing edges, and the spans of the blades extend radially. That is the leading and trailing edges are directed radially.

While I have illustrated specific forms of the invention, it is to be understood that variations may be madetherein and that I intend to claim my invention broadly as indicated by the appended claims.

I claim:

1. In combination in an axial flow compressor for use with an elastic fluid, a case, and an axial flow rotor mounted in said case for rotation about an axis to impel fluid through said case in the general axial direction, said rotor having a hub and a plurality of blades spaced peripherally thereabout at a substantial pitch angle with respect to said axis, said blades and said case defining a plurality of rotor passages whose exit areas and radial depths are substantially greater that the areas and radialdepths of the inlets thereof, the ratio of the diameter of said rotor hub to the tip diameter of said blades in any plane normal to said axis being 0.7 or greater, each said passage having a diffusion angle between the radially inward peripheral surface thereof and said case greater than 12 degrees and less than 45 degrees, the tip section of each said blade from the leading to the trailing edge thereof diverging from a reference line perpendicular to a radius from the rotor axis and parallel to the blade by more than 12 degrees, said passages due to their angle of pitch with respect to said axis being adapted to receive a relative flow thereinto parallel to said blades while providing pumping action on said fluid, and stator blades defining with said case flow passages of decreasing radial extent in the downstream direction for receiving the flow from said rotor.

2. In combination in an axial flow compressor for use with an elastic fluid, a case, and an axial flow rotor mounted in said case for rotation about an axis to impel" fluid through said case in the general axial direction, said case opposite said rotor diverging at an angle greater than 12 and less than 45 with respect to said axis, said rotor having a rotor hub and a plurality of blades spaced peripherally thereabout at a substantial pitch angle with respect to said axis, said blades and said case defining a plurality of rotor passages whose exit areas and radial depths are substantially greater than the areas and radial depths of the inlets thereof, the radial depth of the inlets of said rotor passages being less than the adjacent radius of said hub, said passages due to their angle of pitch with respect to said axis being adapted to receive a relative flow thereinto parallel to said blades while providing pumping and compressing action on said fluid, and axial flow stator passages adjacent the exit side of said rotor adapted to receive the flow of fluid from said rotor, said stator passages having decreasing cross sectional area and radial depth and a length of the order of the maximum radial depth thereof.

3. In combination in an axial flow compressor for use with an elastic fluid, a case, and an axial flow rotor mounted in said case for rotation about an axis to impel fluid through said case in the general axial direction, said case opposite said rotor diverging at an angle greater than 12 and less than 45 with respect to said axis, said rotor having a rotor hub and a plurality of blades spaced peripherally thereabout at a substantial pitch angle with respect to said axis, said blades and said case defining a plurality of rotor passages whose exit areas and radial depths are substantially greater than the areas and radial depths of the inlets thereof, the radial depth of the in lets of said rotor passages being less than the adjacent radius of said hub, said passages due to their angle of pitch with respect to said axis being adapted to receive a relative flow thereinto parallel to said blades while providing pumping and compressing action on said fluid, and axial flow stator passages adjacent the exit side of said rotor adapted to receive the flow of fluid from said rotor, said stator passages having decreasing cross sectional area and radial depth and a length of the order of the maximum radial depth thereof, and a plurality of curved stator blades in said stator ease peripherally disposed therein to define fluid flow passages therebetween, said stator blades having rounded leading edges adjacent the exits of said rotor passages and tapered trailing edges.

4. In combination in an axial flow machine for use with an elastic fluid to raise the static pressure thereof, a case, and an axial flow rotor mounted in said case for rotation about an axis to impel a flow of said fluid therethrough in the general axial direction, said rotor having a hub and a plurality of blades carried thereon closely spaced peripherally thereabout at a substantial pitch angle with respect to said axis, said blades and said case defining a plurality of rotor flow passages whose exit cross sectional areas are in the range of about 1.25 to 3.0 times the corresponding inlet cross sectional areas and whose exit radial depths are substantially greater than the corresponding inlet radial depths, said passages having an axial extent less than the maximum radial length of each said blade, the ratio of the diameter of said rotor hub to the .tip diameter of said blades in any plane normal to said axis being 0.7 or greater, each said passage having a diffusion angle in an axial reference plane between the peripheral surface of said hub and said case greater than 12 degrees and less than 45 degrees providing said range of cross sectional areas within said axial extent of said passages.

References Cited in the file of this patent UNITED STATES PATENTS 461,051 Seymour Oct. 13, 1891 1,447,554 Jones Mar. 6, 1923 1,502,062 Schmidt July 22, 1924 1,614,091 Van Toff Jan. 11, 1927 2,540,968 Thomas Feb. 6, 1951 FOREIGN PATENTS 386,039 Great Britain Jan. 12, 1933 

